Vehicle stabilizer, tilter leveling and pitch control means



L. B. LYNN IL Dec. 20, 1960 2,965,390

VEHICLE STAB IZER, TILTER LEVELING AND PITCH CONTROL MEANS 3 Sheets-Sheet 1 Filed Nov. 5, 1957 INVENTOR Lawrence 8. Lynn Fig.4.

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ATTO NEY Dec. 20, 1960 L. B. LYNN 2,965,390 VEHICLE STABILIZER, TILTER LEVELING AND PITCH CONTROL MEANS Filed Nov. 5, 1957 3 Sheets-Sheet 2 Fig.2.

Dec. 20, 1960 L. B. LYNN 2,965,390

VEHICLE STABILIZER, TILTER LEVELING AND PITCH CONTROL MEANS 3 Sheets-Sheet 3 a .m F

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2,965,390 Patented Dec. 20, 1960 ice VEHICLE STABILIZER, TILTER LEVELING AND PITCH CONTROL MEANS Lawrence B. Lynn, Wilkinsburg, Pa., assignor to Westinghouse Electric Corporation, East Pittsburgh, Pa., a corporation of Pennsylvania Filed Nov. 5, 1957, Ser. No. 694,657 8 Claims. (Cl. 280124) This invention relates to a vehicle stabilizer, tilter, leveling and pitch control means, and more particularly to a vehicle power shock absorber device having a sensing means for stabilizing, tilting and leveling a vehicle sprung mass upon the sensing of vertical, lateral and longitudinal accelerations as well as change in vertical load.

In present day conventional land vehicles, shock absorbing devices located between the sprung mass or body of the vehicle and the unsprung mass (usually comprising vehicle wheels and axles) are of the passive type. This type of shock absorbing device provides reasonable vehicle sprung mass attenuation when the frequency of the unsprung mass movement is above the natural resonant frequency of the sprung mass. However, when the frequency of the unsprung mass movement is approximately the same as the natural resonant frequency of the sprung mass, the passive shock absorber provides an inadequate means of preventing large sprung mass movements.

With the use of passive shock absorbers, for present day vehicles a response to lateral accelerations during a vehicle turn in a manner causing the sprung mass of the vehicle to lean outwardly from the center of the turn is common. Since the sprung mass is usually located above its roll axis, it can be seen that movement of the sprung mass away from the center of the turn tends to cause the center of gravity of the sprung mass to act in a direction to cause over-turning of the vehicle.

It is further common occurrence for a vehicle to respond to longitudinal accelerations or decelerations by a front end rise or dive action and a rear end dip or rising action, respectively.

It is, therefore, an object of this invention to present a vehicle stabilizer providing a means of holding a vehicle longitudinally level during accelerations and decelerations.

It is another object of this invention to provide a sensing device for a power shock absorber that is responsive to longitudinal accelerations.

It is still another object of this invention to provide a power shock absorber system capable of opposing any pitch acceleration.

Other objects, purposes and characteristic features will become obvious as the description of this invention progresses.

It is a practice of this invention to provide stabilizing or shock absorbing means between sprung and unsprung masses at times when the unsprung mass is traversing un even terrain. In the present invention this is accomplished through the use of a plurality of power shock absorbers connected between the sprung and unsprung masses. Each power shock absorber is provided with an acceleration sensing and control means capable of responding to vertical, lateral and longitudinal accelerations in such a way as to cause the sprung mass to appear as a very large equivalent mass and capable of maintaining a constant pitch attitude during accelerations. The sensing and control device of each shock absorber is mounted on the sprung mass and by responding to any acceleration of the sprung mass, either in the vertical, lateral or longitudinal direction, is capable of causing the power shock absorber to rapidly reduce the magnitude of the sprung mass accelerations. The power shock absorbing sensing devices are positioned on the sprung mass in a manner capable of detecting vehicle sprung mass lateral accelerations in response to a turn, and in response to such lateral accelerations causing the power shock absorbers to tilt the sprung mass toward the center of the vehicle circle. The resulting tilt action is continued until the vehicle sprung mass lateral accelerations have become neutralized, at which time the sensing device acts to maintain proper tilt and also to control vertical accelerations due to terrain unevenness.

The sensing devices are also positioned to sense longitudinal accelerations to prevent the sprung mass from pitching due to these accelerations. The present invention also includes a vehicle leveling means utilizing a position controller for maintaining the relative spaced apart positions of the sprung and unsprung masses during a change in vehicle load.

A power driven stabilizer responsive to lateral and vertical accelerations only is shown in the copending application Serial No. 630,835 entitled, Vehicle Stabilizer, Tilter and Leveling Means, filed December 27, 1956 invented by Clinton R. Hanna and assigned to the common assignee.

Figure l is a diagrammatic view of a vehicle frame and under-carriage utilizing four power shock absorbing devices disclosed by this invention;

Fig. 2 is a diagrammatic view of one power shock absorber and its associated power source and sensing elements. The view also includes a diagrammatic showing of a typical sensing element leveling control;

Fig. 2A is a view of the sensing mass and its support pivot springs;

Fig. 3 is a diagrammatic view of one embodiment of sensing device positioning involving this invention;

Fig. 4 is a diagrammatic view of another embodiment of sensing device positioning involving this invention;

Fig. 5 is a graph showing a response curve of the pilot valves in each of the sensing elements in which the pressure is plotted against the valve orifice opening area;

Fig. 6 is a graph showing the relationship of sprung mass shock magnitude to the frequency of the sprung mass movement; and

Fig. 7 is a diagrammatic plan view of still another embodiment of this invention.

In each of the several views similar parts are designated by similar reference characters.

The shock absorbing system shown in Fig. 1 involving four power shock absorbers, to be described in detail hereinafter, is shown applied to a typicalfour wheeled vehicle of a well-known type. The vehicle comprises a frame 1 supported on leaf and coil springs 2 which are in turn secured to axles and wheels 3 in a conventional nature. The axles and wheels 3 will be referred to hereinafter as the unsprung mass of the vehicle, while the frame member 1 and its associated parts, such as an engine and a typical body (not shown) will be referred to as the sprung mass. It is pointed out at this time that the sprung mass is considerably heavier than the unsprung mass which is the typical situation with present day land vehicles. This means that the natural frequency of the sprung mass is considerably below that of the unsprung mass, which is also a desirable characteristic when attempting to reduce sprung mass oscillations.

The power shock absorbing system shown in Fig. I is provided with a pump 5 for supplying fluid under pressure and a pair of accumulators 6 and 7. The pump 5 also includes a return reservoir such as the reservoir 8, shown in Fig. 2. The pump 5 has snfiicient capacity to provide fluid under pressure for all four of the power shock absorbers as well as providing fluid under pressure for other auxiliary equipment (not shown) Within the vehicle. The two accumulators 6 and 7 are placed'at opposite ends of the vehicle in order to provide reasonably constant fluid pressures to the nearest pair of power shock absorbers. In addition, the pump is provided with an unloading valve 9, shown only diagrammatically in Fig. 2, for bypassing or circulating fluid pumped by the pump 5 after the desired pressure level has been reached in each of the accumulators and during little or no shock absorbing action.

The view of Fig. 1 discloses the use of four power shock absorbers, generally designated as 11, 12, 13 and 14. Each power shock absorber is provided with sensing and control devices, designated 15, 16, 17 and 18, respectively. Since each of the power shock absorbers performs in an identical manner, a typical power shock absorber is shown in Fig. 4 and is used to describe the operation of any one of the power shock absorbers 11, 12, 13 and 14.

If we assume that the typical shock absorber, shown in Fig. 2 is the shock absorber designated generally as 11 positioned on the left of the vehicle as viewed from the rear of Fig. 1, it can be seen that the structure has a sprung mass 1 secured by a bushing 19 to an upper portion 19a housing control valves and a lower portion or piston rod 20 which is secured to the unsprung mass 3 by a bushing 21 (see Fig. 2). The upper portion 19a, secured to the sprung mass 1, is provided with an integral cylinder 22 extending downwardly for receiving a piston 23 which is integral with the lower portion or piston rod 20. The piston rod 20 is equal to one half of the downwardly faced area of the piston 23 and is provided with a fluid sealing slip bushing 24 secured to the cylinder 22 of the upper portion 19. The bushing 24 is provided with a fluid annular ring leakage return path 24a. for preventing fluid leakage from the high pressure area within the cylinder 22 to the low pressure of the outside area. It also serves to reduce the friction of bushing 24 which would be quite high if subjected to the average cylinder pressure. The piston 23 and the end sealing bushing 24 in cooperation with the cylinder 22 form a chamber 25 provided with fluid under pressure from the associated accumulator 7 delivered through the pipe 26 in a manner to be described hereinafter. A second chamber 27 is formed by a cylinder 22 and the end of the piston 23 opposite to the piston rod 20. The chamber 27 is connected to a pair of fluid control valves 28 and 29 so that fluid may be admitted to or relieved from the chamber 27 in a manner demanded by the sensing element and the movement of the piston 23 to be described hereinafter. The valve 28 is a pressure fluid admission valve capable of delivering fluid from the accumulator 7 through the supply pipe 30, past the valve closure end 31 into the chamber 27. The valve 28 is provided with a piston 32, subjected on one side to the accumulator pressure and on the other side to a balancing force pressure determined by the sensing element, as explained hereinafter.

The valve 29 is an exhaust valve capable of allowing fluid to flow from the chamber 27, out past its valve seat 33 and into the fluid return pipe 34. The valve seat 33 of the exhaust valve 29 is a spring biased seat which is normally stationary but operable under emergency conditions of power failure to prevent a completely inoperative shock absorber, as explained hereinafter. It is sufficient to note at this time that upon a pressure drop in the chamber 27 during fluid supply failure, the valve seat 33 is drawn downwardly causing an associated coil bias spring 35 to be compressed. This allows fluid to flow from the return pipe 34, past the valve seat 33 and into the chamber 27. The exhaust valve 29, similar to the intake valve 28, is also provided with an actuating piston 36 which is subjected to accumulator pressure on the valve side of the piston and to a lower sensing element adjusted pressure on the side of the piston opposite the valve seat 33.

Accumulator pressure acting against the valve closure end 31 and the valve closure end side of the piston 32 results in a force in a direction to seat the valve end as shown in Fig. 2. This force is opposed by the sensing element 15 varied fluid pressure received through a pipe 37 to act on the larger area of the opposite side of the piston 32. Under normal conditions, with no accelerations being detected by the sensing element or device 15, the forces on the piston 32 of the intake valve 28 are approximately equal with the valve 31 being seated by the pressure applied by the fluid in chamber 27. Likewise the pressures supplied to the piston 36 by the accumulator and the sensing feed pipe 37 as well as the chamber 27 applied pressure on the valve, all result in approximately no control or moving forces with the exhaust valve holding in its seated position.

The above-described relationship of pressure to piston areas of the valves 28 and 29 is established so that the response of the valves is as near to a linear response as possible. This is shown in the pressure pilot orifice curve of Fig. 5 described in greater detail hereinafter. With the valve forces so equalized, movement of the shock absorber piston 23 can command immediate fluid movement into or out of the chamber 27 with the sensing and control device then taking control of the pressure and exhaust valves 28 and 29 in response to sprung mass movement.

Located within the pressure feed from the accumulator '7 to the lower chamber 25 of the power shock absorber 11 is a pair of damper valves 38 and 39. The damper valves 38 and 39 are positioned in fluid pipes 40 and 41, respectively, interconnecting the main fluid supply pipe 3% with the lower chamber 25 through the feeder pipe 26. The damper valve 38 is a spring controlled valve capable of allowing flow from the chamber 25 to the accumulator 7 only. This valve comprises a movable member 42 held to its seated position by a spring 4-3 with zero force for that position.

The damper valve 39 is a similar valve positioned to allow flow from the accumulator 7 to the lower chamber 25 only. This valve involves a movable member 44 provided with a spring 45 for seating the valve in the position shown. Under the conditions of inactivity shown in Fig. 2, the pressure within the chamber 25 is equal to the accumulator 7 pressure.

The movable member 44 is provided with an orifice 44a passing therethrough for causing the action of the damper valve 39, as well as the damper valve 38, to follow a more nearly straight line of response. The orifice 44a is needed since without the orifice the rate of fluid transfer past the damper valves is slow at the beginning of a difference in pressure and increases as the pressure rises. The action of the orifice 44a to a difference in pressure is the opposite, with a rapid flow rate of change at the beginning and a reduced rate-of change as the pressure rises. The two actions combined thus result in a more nearly straight line response curve.

The piston 23 within the cylinder 22 is provided with a normally closed emergency flow valve 46 having a movable member 46a biased to its closed position by a bias spring 47. The valve 46 is used to interconnect the chamber 25 with the chamber 27 under conditions of power source failure, as will be explained hereinafter.

The power shock absorber is controlled by a sensing and control device 15, which, as explained hereinbefore, is securely fastened to the frame or the sprung mass 1 of. the vehicle. The sensing and control device 15 is secured to the frame at an angle as shown in Figs. 3, 4 and 7 and comprises a weight 48 supported on a pair of leaf springs 49 (see Fig. 2A), adjacent one edge such that the center of gravity of the weight 48 is to one side and above the leaf spring supports 49. Projecting outwardly from the weight 48, adjacent to and on the opposite side of the support springs 49, is a pilot valve control arm 50. Positioned adjacent to and slightly spaced from the pilot valve control arm 50 is a pilot valve thimble 51, having an orifice 52 in the end thereof. The pilot valve thimble 51 is connected to the accumulator 7 through an orifice 53, filter 54 and the pressure pipes 55 and 30. The pilot valve thimble is also connected to the intake and exhaust valves 28 and 29 of the power shock absorber 11 through the previously mentioned pilot valve control pipe 37. It can be seen, therefore, that the accumulator fluid, under pressure, is fed through the pipe 30, the pipe 55, the filter 54 and orifice 53 filling the pilot valve control fluid pipe 37 and the pilot valve thimble 51 and flowing through the orifice 52 in the thimble against the pilot valve control arm 50. The fluid flow out of the orifice 52 fills the chamber about the sensing weight 48 within the cover 56 and returns to the reservoir 8 through a return pipe 57. It can be seen that controlled fluid flow occurs through the orifice 52 with the amount of flow being determined by the proximity of the pilot valve control arm 50 to the orifice 52. This flow variation through the orifice 52 in turn causes a pressure variation on the control pistons of the power shock absorber intake and exhaust valves 28 and 29, respectively.

Since the pivot point of the weight 48 is to one side of the center of gravity, it is necessary to provide a static balance for the weight 48. This static balance is provided through a piston 58 positioned within an opening 59 in the base 60 of the sensing and control device 15. The piston 58 is biased upwardly against the bottom of the weight 48 by a bias spring 61 provided with a variable support 62 for varying the amount of bias presented to the sensing and control weight 48. In addition, the piston 58 is provided with fluid damping through a needle valve 63 having an orifice and pipe 64 interconnecting the area within the sensing weight 48 cover or shell 56 with the area below the piston 58, housing the 'bias spring 61. In this way, fluid enters past the needle valve 63 through the passage 64, and is restricted in its movement into and out of the area below the piston 58 by the size of the orifice which is adjusted by the needle valve 63. The action of this fluid damping is necessary to eliminate resonant vibrations of the control mass.

Since variations in the accumulator pressure supplied by the pump 5 would inherently affect the balance of the weight 48, a pressure compensating piston 65 is positioned in contact with the bottom of the sensing weight 48. The compensating piston 65 is positioned within a bore 66 connected with the filter chamber 54 and fluid supply pipe 55. It can be seen then that variations in the accumulator pressure will then be reflected in the force of piston 65 on the sensing weight 48 to compensate the average forces of the orifice 52 and piston 67 on the opposite side of the pivot 49. Thus the weight 48 assumes the same average position for varying accumulator pressures.

In order to introduce velocity control to the sensing weight 48 in addition to the acceleration control previously described, a feedback piston 67 is provided. The piston 67 cooperates with the pilot valve control arm 50 at a point to the left of the pivot springs 49, as viewed in Fig. 2. The feedback piston is housed within a bore 68 in the base 60 and is connected to the chamber 25 of the power shock absorber by a fluid control pipe 69. For this reason the response of the feedback piston 67 is proportional to the relative velocity of the movement of the piston 23 within the cylinder 22 reflected in relatively small pressure changes in the chamber 25 above or below the accumulator pressure during fluid flow through one of the damper valves 38 or 39.

The sensitivity of the sensing mass 48 to vertical and lateral forces can be varied by two major methods.

The first method, not used in this system, is to vary 6 the shape of the sensing mass to cause the center of gravity to move closer to or farther away from the pivot point of the sensing mass. This system obviously has the disadvantage of requiring a modification of the sensing mass which is difiicult.

The second method is to vary the position of the sensing mass with respect to its pivot point causing a change in the effective lever arm length through which the forces can act. This is the method of control used in this application and is readily and rapidly changed or varied as desired.

Operation of the power shock absorber system of Fig. 2, as described up to this point, will now be explained. If we assume that the vehicle is passing over an uneven surface in which the unsprung mass 3 is driven upwardly toward the sprung mass of the frame 1, the action of this motion through the power shock absorber and the supporting springs 2 would then cause a vertical acceleration on the weight 48 of the sensing device 15. Since the motion of the frame or the sprung mass 1 is in an upward direction, the weight 48 resists the upward movement resulting in a lagging of the weight downwardly causing the pilot valve orifice 52 to be relieved, allowing greater fluid flow through the orifice. This is due to the weight 48 pivoting clockwise about the leaf springs 49, since the center of gravity of the weight is to the right of the pivot springs 49.

An increase in flow through the orifice 52 results in a reduction of pressure in the pilot control pipe 37 and on the pistons 32 and 36 of the intake and exhaust valves 28 and 29 of the power shock absorber. The result in reduction in pressure above each of the pistons 32 and 36 then causes the accumulator pressure on the under side of the pistons 32 and 36 to result in a more firm seating of the intake valve 28 and an opening of the exhaust valve 29. When this occurs, a rapid reduction in pressure within the chamber 27 of the power shock absorber takes place, allowing the piston 23 to move rapidly upward in response to the uneven terrain. The pressure from the accumulator in chamber 25 causm the force on the under side of the piston 23 to exceed the force on the upper side with the result that upward accelerations of the sprung mass are greatly impeded. The rapid upward movement of the piston 23 reduces the pressure within the chamber 25 slightly, however, causing the damper valve 39 to open and supply more fluid to the chamber 25. The reduction in pressure within the chamber 25 also results in a reduction of the force imposed on the pilot valve control arm 50 by the feedback piston 67. This action tends to reduce the flow through the pilot valve orifice 52 with the result that damping of the resonances of the vehicle masses occurs.

The graph shown in Fig. 6 discloses a family of curves showing the response of the sprung mass of a vehicle with different types of shock absorbers applied thereto, and with the unsprung mass being moved at constant amplitude over a range of frequencies from less than one cycle per second to six cycles per second.

Curve A is a curve showing the sprung mass response with the power shock absorber of this invention applied thereto. As shown by this curve, the peak shock rise of the spring mass occurs at a point well below one cycle per second and is only 1.2 units in amplitude.

Curve B is a curve showing the sprung mass response using a conventional passive type shock absorber of heavy damping quality. The peak shock rise with this device occurs at approximately the one cycle per second period and is nearly 1.7 units in amplitude. In addition, the magnitude of the response of the sprung mass at higher frequencies is greater.

In order to reduce the magnitude of the high frequency response of the sprung mass while using passive type shock absorbers, the damping action of the shock absorber is reduced as shown by curve C. It is obvious, however,

7 that in order to obtain good high frequency response with passive type shock absorbers a sacrifice in low frequency response occurs. As shown by curve C the peak rise occurs at the one cycle per second frequency and is 2.1 units in amplitude.

From the results shown by the family of curves of Fig. 6 it is clear that control of pressures within the power shock absorber of this invention through the use of an acceleration sensing device provides greatly improved sprung mass response as compared to response under conventional passive shock absorber control devices.

Referring again to Figure 2 under operation involving a separation of the sprung and unsprung masses due to the unsprung mass 3 following a depression in the surface being traversed, it can be seen that a pressure rise within the chamber 25 takes place. The increase in pressure within the chamber 25 as the piston 23 attempts to move downward along with the unsprung mass causes the sprung mass to attempt a following action. The acceleration of the sprung mass downwardly causes the weight 48 to pivot counterclockwise about its support springs 49, resulting in a restriction of the orifice 52 in the pilot thimble 51. This restriction causes a rise in pressure in the chambers above the pistons 32 and 36 of the intake and exhaust valves, respectively. The rise in pressure above each of the pistons 32 and 36 results in an opening of the intake valve 23 and a more firm seating of the exhaust valve 29. Opening of the intake valve 28 results in an increase in pressure within the chamber 27, aiding the piston 23 in its down ward movement due to the separation of the sprung and unsprung masses. The resultant increase in pressure within the chamber 25 is then relieved through the damper valve 42 into the accumulator 7 system. This rise in pressure within the chamber 25 also results in a feedback pressure increase on the piston 67. The action of the piston 67 then results in a clockwise rotation tendency of the weight 48 with the end result being a velocity stabilization of the original or initial action of the weight 48.

It should be pointed out at this time that the pressure within the chamber 25 is approximately twice that of the pressure within the chamber 27 under periods of stabilization. The resultant forces on the piston 23 are equal and opposite since the affected area of the piston 23 in the high pressure zone of the chamber 25 is one-half of the area of the piston 23 exposed to the chamber 27. For example, a typical accumulator pressure might be on the order of 2000 p.s.i. with the resulting pressure within the chamber 25 under static conditions also being 2000 p.s.i. The pressure within the chamber 27, however, would only be 1000 p.s.i. due to the action of the valves 28 and 29. This, theoretically, provides an active range for the control valves 28 and 29 of from zero to 2000 .s.i. within the chamber 27. The pressures of the chamber 25 vary a smaller amount, however, this amount occurs about the 2000 p.s.i. level. It is this difference in pressure between the chamber 25 and the chamber 27 during operating of the pump supplying pressure to the power shock absorber that maintains the emergency flow valve 46 closed at all times.

It should also be pointed out that the intake and exhaust valves 28 and 29 are operated at a high pilot pres sure level in order to have the action of these valves occur over the most nearly straight portion of the pressure response curve for the valves, as shown in Fig. 5. As shown by this graph, plotted for a 2000 p.s.i. accumulator pressure, the straight line portion of the curve occurs between maximum and minimum pilot valve orifice opening establishing pilot pressures of between 1000 p.s.i. and 2000 p.s.i. The piston area relationships of the valves 28 and 29 are thus selected to allow valve operation within this pressure zone.

Operation of the power shock absorber in response to lateral accelerations will now be explained. If we assume the vehicle is starting into a turn to the right, as

viewed in Fig. 2, with the assumption that the power shock absorber is located on the left side of the vehicle as seen from the rear, such as the power shock absorber 11 of Fig. 1, the acceleration of the vehicle to the right causes the weight 48 to pivot counterclockwise about the leaf springs 49. This results in a restriction of the orifice 52 of the pilot valve thimble 51, resulting in a rise in pressure in the areas above the pistons 32 and 36 of the feed and exhaust valves 28 and 29, respectively. As explained hereinbefore, this results in an increase in pressure above the pistons 32 and 36 causing the intake valve 28 to open and the exhaust valve 29 to become more firmly seated. The opening of the valve 23 results in an increase in pressure within the chamber 27 and a movement of the piston 23 downwardly, separating the sprung and unsprung masses. Since this power shock absorber is indicated as being located on the left side of the vehicle as seen from the rear, it should be obvious that a tilting action of the vehicle sprung mass toward the center of the circle takes place.

At the same time, the power shock absorber 12 on the right side of the vehicle, as seen from the rear of the vehicle of Fig. l, with its sensing element 16 positioned approximately with respect to the sensing element 15 of the shock absorber 11, would cause the sprung and unsprung masses to be drawn together. If we assume that the sensing element 15 of Fig. 2 is reversed approximately 180 from the position shown therein, it can be seen that the weight 48 would be pivoted by its support leaf springs 49 in a direction to cause a relief of the orifice 52, allowing greater flow from the orifice and a resulting in a reduction in pressure within the chambers above pistons 32 and 36 of the valves 28 and 29, respectively. As explained hereinbefore, the reduction in pressure causes the exhaust valve 29 to open and the intake valve 28 to be more fully seated. This results in a reduction in pressure within the chamber 27, causing the piston 23 to move upwardly within the cylinder 22. The result is a reduction of the distance or spacing between the sprung and unsprung masses on the right side of the vehicle.

Since the piston of the power shock absorber 14 and its sensing element 18 is similar to that of the shock absorber 11, that is, being positioned on the left side of the vehicle, it can be seen that the previously mentioned turn to the right would also cause an extension of the power shock absorber 14 in a manner similar to the power shock absorber 11. Likewise, the power shock absorber 13 would have a resulting action to a vehicle turn to the right that is similar to the power shock absorber 12, that is, a contraction of the power shock absorber elements, causing the space between the sprung and unsprung masses to be reduced as the masses are pulled toward each other on the right side of the vehicle.

The cooperating action of all four shock absorbers shown in Fig. l in response to a turn to the right, would therefore result in a tilting of the sprung mass, involving the frame 1 of the vehicle, toward the center of the circle.

It is pointed out that with the present invention the acceleration response of the power stabilizer effectively increases the equivalent sprung mass of the vehicle several fold. Any force, such as the vehicle spring force or damping force due to high velocity of the unsprung mass relative to the sprung mass, therefore causes much lower acceleration of the body than with conventional shock absorbers. Such lower accelerations are the same as would occur had the springs been softer and the relative damp ing lower. It can be shown that insofar as the unsprung mass movements are concerned, the damping is in actuality reduced.

Let the equivalent mass M due to the acceleration response be related to the body mass m. by a factor K.

For critical damping of the body resonance with springs having stiffness S, the damping is if K=8, then R=6 /mS, which value is much higher than is tolerable with conventional shock absorbers.

The available damping for the unsprung mass can be shown to be For K=8 as before, then r= /a /;1:S, a smaller but reasonable value which is of the order needed for wheel damping.

The combination of high acceleration response with high damping due to relative movement, therefore, provides greatly increased impedance to body movements but normal low damping impedance to wheel movements.

At times of power failure it is necessary for the power shock absorbers to provide passive shock absorber action similar to conventional shock absorbers. The previously described movable valve seat 33 and emergency valve 46 provides the means of satisfactory emergency opera tion.

If the unsprung mass 3 causes a downward movement of the piston 23 during power failure, the chamber 27 experiences drop in pressure causing a fluid demand. This demand is satisfied by the opening of the valve seat 33 against its bias spring 35, allowing fluid flow from the return pipe 34 into the chamber 27. At the same time the chamber 25 experiences a rise in pressure which is allowed to escape through the damper valve 38 into the accumulator 7 system.

Upon a movement of the unsprung mass 3 and piston 23 upwardly, the pressure within the chamber 27 rises. With impeded fluid escape through the valves 28 and 29, the fluid is relieved through the emergency valve 46 into the chamber 25 which is simultaneously experiencing a reduction of pressure.

The movement of fluid through the valves 38 and 46 and the valve seat 33 provides adequate passive shock absorber action. This passive shock absorber action may actually be improved due to the partial inert-la control which can occur as the accumulator is charged by the pumping action of the chamber 25.

Operation of the power shock absorber in response to longitudinal acceleration of the sprung mass utilizing the sensing devices positioned at an angle to the lateral axis of the sprung mass as shown in Figure 3 will now be explained.

If we assume that the vehicle is traveling at a desired velocity and is brought to a stop through the use of its brakes, a diving action of the front end of the sprung mass as well as a rising action of the rear end of the sprung mass takes place. This action may be referred to as a pitch motion and occurs due to the fact that the center of gravity of the sprung mass is located above the support for the sprung mass. In order to prevent the diving action of the front end of the vehicle and the rising action of the rear end of the vehicle, the sensing and control devices of each are turned through an angle about a vertical axis through the centering device. Rotation of the sensing device and consequently the sensing mass 48 about a vertical axis passing through the pivot springs 49 provides an effective lever arm along the longitudinal axis of the sprung mass of the vehicle. The length of the longitudinal lever arm is varied by the angular position of the sensing device with respect to the logintudinal axis Of the sprung mass. The sensing devices are rotated about the vertical axes of the pivots 49 in a direction necessary to allow the sensing masses 48 to deliver fluid under pressure to its associated front shock absorbers in response to movement of the masses 48 upon deceleration of the sprung mass supporting the sensing masses 48.

If we consider that the absorber of Fig. 2 is the absorber 14 positioned on the left side of the vehicle as shown in Figures 1 and 3 and with the sensing device 18 placed at an angle as shown in Figure 3, with the mass 48 located rearward of the pivot point 49, a restriction of the orifice 52 occurs during braking action causing the mass 48 to pivot to the left, as viewed in Fig. 2 about the springs 49. Restriction of the orifice 52 causes an increase in pressure in the cylinder 28 above the intake valve 31, opening the intake valve 31 to admit pressure into the shock absorber upper chamber 27 causing it to extend. The action of the shock absorber, therefore, is in opposition to the decelerating forces applied to the sprung mass of the vehicle. The amount of force desired to be delivered by the shock absorber for the purpose of maintaining the vehicle sprung mass level during deceleration is determined by the length of the lever arm along the longitudinal axis of the vehicle through which the sensing mass 48 acts. This length is varied through theangle. The sensing device is rotated in the initial setting of the device in the sprung mass. At the same time, the shock absorber 13 and sensing device 17 are performing a similar function aiding the previously mentioned shock absorber 14 and sensing device 18 as described above.

During this braking action there is also a tendency for the rear end of the Sprung mass to rise. In order to compensate for this rising tendency of the rear end of the vehicle, the sensing device 15 and 16 are rotated about their pivot springs, such as a pivot spring 49 of Fig. 2, in a direction opposite to the previously described front sensing mass devices. That is, the sensing masses such as the sensing mass 48 of Fig. 2 is moved forward of the pivot point 49 a sufficient distance to provide a lever arm for the sensing mass 48 along the longitudinal axis of the sprung mass. During the deceleration period, it can be seen that the mass 48 would respond by .a rotation to the right. as viewed in Fig. 2. when placed on the rear of the vehicle with the mass forward of the pivot spring 49. This action causes a reduction in pressure above the pistons 32 and 36 of the intake and exhaust valves 31 and 33, respectively, causing the exhaust valve to open and the intake valve to remain seated. With a reduction in pressure from the chamber 27 above the piston 23 of the shock absorber such as shown in Fig. 2, a shortening of the shock absorber takes place. This action serves to hold the rear end of the vehicle from rising in response to the deceleration.

Since rotation of the sensing device about their support springs 49 affects the length of the effective lever arm previously described in connection with lateral control of this vehicle sprung mass and since the lever arm for vertical action of the sensing mass is unchanged, it can be seen that the longitudinal sensing becomes the product of the lateral gain of each of the sensing devices times the angular displacement of the sensing mass.

If it is desired to provide the longitudinal brake dive control as a function of the vertical gain of the sensing device, it is merely necessary to tilt the sensing device about a horizontal lateral axis a suflicient amount to cause the center of gravity of the sensing masses to fall forward of or to the rear of the pivot point 49, as shown in Fig. 4. With the sensing mass of each sensing de ice positioned across the vehicle along a transverse axis of the sprung mass, it can be seen that any acceleration or deceleration of the sprung mass along the longitudinal axis of the sprung mass will not efiect the gain of the lateral response of the sensing mass of each device. With each device tilted, however, placing the center of gravity of each sensing mass forward or rearward of the pivot point, an effective lever arm is produced causing the sensing mass to respond to longitudinal acceleration as a 11 function of the vertical response of the sensing mass 48.

Under certain circumstances, it may be desirable to provide the longitudinal control as a function of both lateral and vertical gain of the sensing mass 48 of each of the sensing devices on the vehicle. This arrangement is shown in Fig. 7 where the sensing mass is displaced at angles in both the vertical and horizontal planes of the vehicle sprung mass. In this case, the sensing mass has an effective lever arm that reduces slightly for small displacement angles the gain of the vertical and lateral controls of the sensing mass 48. The brake dive control then becomes the sum of the gain of the vertical and lateral responses, times their respective displacement angles.

It should be pointed out that in each case of the positioning of the sensing device shown in Figs. 3, 4, and 7 explained as opposing brake dive characteristics or deceleration characteristics serve equally as well in the control of front end rise and rear end dip occurring during acceleration of the vehicle.

The angle at which the sensing device must be placed in order to detect longitudinal acceleration or' deceleration can be expressed mathematically by the following equations.

The magnitude of the pitch disturbance torque can be expressed by:

where T =hM a T =disturbance torque h=separation of gravity and pitch centers M =sprung mass a=longitudinal acceleration If we assume for the purpose of illustration that the vehicle is symmetrical and that the pitch center and the center of gravity lie on a vertical centerline through the sprung mass such as shown in Fig. 3, and if we assume that the stabilizer component characteristics are identical, then the corrective torque can be explained as follows:

where l=wheel base =lateral gain factor M =sensing element mass =longitudinal acceleration =angle between pivot axes and longitudinal vehicle axis To provide exact pitch torque cancellation, the corrective torque must be equal to the disturbance torque, therefore, the two previously presented equations must be equal.

Equating these equations results in the following equation when reduced:

(3) sin 0= 4 ,M,=sM, or gr y.

from which If this equation is used for determining the angle of Sin 6 "the sensing devices ofFig. 4, the same formulas hold 12 true with the exception that the gain factor for the lateral response is replaced by the gain factor for the vertical response of the stabilizer system.

If the sensing elements are displaced in both planes, the following relationship should obtain:

The system of Fig. 2 also discloses a leveling control for the shock absorber 11. This leveling control is also used to control the shock absorber 12 of Fig. l. The purpose of the leveling control is to provide a readjustment of the relative distances between the sprung and unsprung masses when a change in load occurs within the sprung mass 1 including the vehicle body (not shown). The leveling structure shown in Fig. 2 comprises a leveling control switch 70 having two pairs of contacts 70a and 70b supported on the sprung mass of the vehicle. Associated with the two pairs of contacts 70a and 70b is a shunting bar 700 capable of selectively bridging each pair of contacts upon certain changes of load conditions within the vehicle. The bridging contact 70c is controlled by a resilient leaf spring 71 having one end secured to the under-carriage or unsprung mass 3 and the other end secured to the spring mass 1, as shown in Fig. 2. The leaf spring 71 provides a means of introducing a large motion ratio between the unsprung mass 3 movement and bridging contact 70c movement. The addition of a load within the vehicle causing the sprung mass 1 to settle causes the bridging contact 70c to bridge the pair of contacts 70a completing an electrical circuit, to be described hereinafter. If, however, a load is removed from the vehicle, causing the sprung mass 1 to rise, with respect to the unsprung mass, the bridging contact 700 is caused to bridge the pair of contacts 70b completing a second electrical circuit, to be described hereinafter.

Supported on the sprung mass 1 in a position to reposition the ,variable support 62 of the bias piston 58 in the sensing element 15, is a gear train 72 having a gear ratio input to output of several hundred to one. The output of the gear train 72 is delivered to the variable support 62 through a vertically moving shaft 73. The gear train 72 is provided with an input shaft 74 driven by a reversible motor 75. As shown in Fig. 4, the motor 75 is provided with a pair of windings 76 and 77 for causing the motor to rotate'in one direction or the other.

The switch 70 has one of each of its pairs of contacts 70a and 70b connected in parallel and through a source of power 78, a conductor 79 to the armature of the motor 75. The other terminal of the pairs of contacts, indicated as 70a, is then connected through a conductor 80 to the winding 76, which has its other end connected to the armature of the motor 75. Likewise, the other contact of the pairs of contacts 70b is connected through a conductor 81 to the winding 77, which is also connected to the armature of the motor 75.

It can be seen, therefore, that displacement of the sprung mass downwardly causes the bridging contact 700 to bridge the contacts 70a, energizing the motor 75, through its winding 76, causing a rotation of the shaft 74 and an extension of the output shaft 73 of the gear train 72. Extension of the shaft 73 in turn increases the bias pressure on the piston 58, which then causes the sensing weight 48 to pivot counterclockwise about its resilient leaf springs 49, restricting the output of the pilot orifice 52 and thus increasing the fluid fiowand pressure within the chamber 27 of the power shock absorber 11, as explained hereinbefore. This action drives the piston 23 downwardly causing separation movement of the sprung and unsprung masses until the contact 700 is again separated from the contacts 70a.

If a load is removed from the vehicle sprung mass, it can be seen that the contact 70c then bridges the contacts 70b, causing a reverseaction to take place. It should be 13 pointed out at this time that it is necessary for a time delay to be introduced between the action of the contact switch 70 and the resulting motion delivered by the power shock absorber 11. This time delay is desirably from one to two minutes in length and it is introduced by the high ratio of the gear train 72.

Since numerous changes may be made in the abovedescribed construction and different embodiments of the invention may be made without departing from the spirit and scope thereof, it is intended that all the matter contained in the foregoing description or shown in the accompanying drawings shall be interpreted as illustrative and not in a limiting sense.

I claim as my invention:

1. In a stabilizer for a vehicle having two relatively moving masses, power driven extensible force producing means connected between said masses, acceleration sensing and control means, said acceleration sensing and control means including a plurality of sensing masses, pivot means for each said plurality of sensing masses for securing said sensing masses to one of said moving masses, each said sensing mass being pivoted at a position remote from its center of gravity with the pivot positioned to provide effective lever arms along the lateral, vertical and horizontal axis of said one moving mass to thus, by the pivotal movement of said sensing masses, control said force producing means.

2. In a stabilizer for a vehicle having two relatively moving masses, power driven extensible force producing means connected between said masses, acceleration sensing and control means, said acceleration sensing and control means including a plurality of sensing masses, pivot means for each said plurality of sensing masses for securing said sensing masses to one of said moving masses, each said sensing mass being pivoted on a hinge axis at a position remote from its center of gravity with the pivot axis positioned to provide effective lever arms along the lateral, vertical and horizontal axis of one of said moving masses to thus, by the pivotal movement of said sensing masses, control said force producing means, said hinge axis being placed at an angle to the lateral and longitudinal axis of one of said moving masses to thus provide an effective longitudinal lever arm action as a function of movement of sensing mass lateral axis control.

3. In a stabilizer for a vehicle having two relatively moving masses, power driven extensible force producing means connected between said masses, acceleration sensing and control means, said acceleration sensing and control means comprising a plurality of sensing masses, pivot means for each said plurality of sensing masses for securing said sensing masses to one of said moving masses, each said sensing mass being pivoted on a hinge axis at a position remote from its center of gravity with the pivot axis positioned to provide effective lever arms along the lateral, vertical and horizontal axis one of said moving masses to thus, by the pivotal movement of said sensing masses, control said force producing means, said hinge axis being at an angle to the longitudinal axis and perpendicular to the lateral axis of one of said moving masses to thus provide an effective longitudinal lever arm action as a function of movement of sensing mass vertical response control.

4. In a stabilizer for a vehicle having two relatively moving masses, power driven extensible force producing means connected between said masses, a sensing and control means for controlling said force producing means in response to longitudinal accelerations of one of said masses, said sensing and control means being mounted on said one of said masses, said sensing and control means also producing forces in said power producing means connected between said masses in response to the relative velocities of said masses, said sensing and control means having a sensing mass pivoted to one side and below the center of gravity of said sensing mass, said sensing means being secured to said one mass at an angle to the longitudinal axis of said one mass and remote to the center of gravity of said one mass to provide a longitudinal effective lever arm for the sensing mass.

5. In a stabilizer for a vehicle having two relatively moving masses, power driven extensible force producing means connected between said masses, a sensing and control means for controlling said force producing means in response to longitudinal accelerations of one of said masses, said sensing and control means being mounted on said one of said masses, said sensing and control means also producing forces in said power producing means connected between said masses in response to the relative velocities of said masses, said sensing and control means having a sensing mass pivoted to one side and below the center of gravity of said sensing mass, said sensing means being secured to said one mass at an angle to the longitudinal axis of said one mass and remote to the center of gravity of said one mass to provide a longitudinal effective lever arm for the sensing mass, the angle position of said sensing mass with respect to said longitudinal axis being such that its control is to increase the extensible force produced by said force producing means during deceleration periods when the sensing mass is placed forward of the center of gravity of said one mass.

6. In a stabilizer for a vehicle having two relatively moving masses, power driven extensible force producing means connected between said masses, a sensing and control means for controlling said force producing means in response to longitudinal accelerations of one of said masses, said sensing and control means being mounted on said one of said masses, said sensing and control means also producing forces in said power producing means connected between said masses in response to the relative velocities of said masses, said sensing and control means having a sensing mass pivoted to one side and below the center of gravity of said sensing mass, said sensing means being secured to said one mass at an angle to the longitudinal axis of said one mass and remote to the center of gravity of said one mass to provide a longitudinal effective lever arm for the sensing mass, the angle position of said sensing mass with respect to said longitudinal axis being such that its control is to increase the extensible force produced by said force producing means during deceleration periods when the sensing mass is placed forward of the center of gravity of said one mass, and is such that the sensing mass control acts to decrease the extensible force produced by said force producing means during deceleration periods when the sensing mass is placed rearward of the center of gravity of said one mass.

7. In a stabilizer for a vehicle having two relatively moving masses, power driven extensible force producing means connected between said masses, a sensing and control means for controlling said force producing means in response to longitudinal accelerations of one of said masses, said sensing and control means being mounted on said one of said masses, said sensing and control means also producing forces in said power producing means connected between said masses in response to the relative velocities of said masses, said sensing and control means having a sensing mass pivoted to one side and below the center of gravity of said sensing mass, said sensing means being secured to said one mass at an angle to the longitudinal axis of said one mass and remote from the center of gravity of said one mass to provide a longitudinal effective lever arm for the sensing mass, the angle position of said sensing mass with respect to said longitudinal axis being such that its control is to decrease the extensible force produced by said force producing means during acceleration periods when the sensing mass is placed forward of the center of gravity of said one mass.

8. In a stabilizer for a vehicle having two relatively moving masses, power driven extensible force producing means connected between said masses, a sensing and 15 control means for controlling said force producing means in response to longitudinal accelerations of one of said masses, said sensing and control means being mounted on said one of said masses, said sensing and control means also producing forces in said power producing means connected between said masses in response to the relative velocities of said masses, said sensing and control means having a sensing mass pivoted to one side and below the center of gravity of said sensing mass, said sensing means being secured to said one mass at an angle to the longitudinal axis of said one mass and remote from the center of gravity of said one mass to provide a longitudinal efiective lever arm for the sensing mass, the angle position of said sensing mass with respect to said longitudinal axis being such that its control is to References Cited in the file of this patent UNITED STATES PATENTS 2,492,990 Hanna Jan. 3, 1950 2,593,040 Lloyd Apr. 15, 1952 2,657,067 Kolbe Oct. 27, 1953 2,778,656

May Jan. 22, 1957 

